KR20210022930A - Non-azeotropic mixed refrigerant, and refrigerating apparatus using the same - Google Patents

Abstract

According to the present invention, a refrigerating device using a non-azeotropic refrigerant mixture comprises: a compressor operated in a continuous driving mode to compress a non-azeotropic refrigerant mixture; a condenser condensing the refrigerant compressed in the compressor; an expander expanding the refrigerant condensed in the condenser; and an evaporator vaporizing the refrigerant expanded in the expander. A pressure difference (ΔP) of the non-azeotropic refrigerant mixture has a value which falls within 340 kPa < ΔP < 624.7 kPa. Accordingly, reliability of a driving component such as a piston and the like can be further improved in the refrigerating device using the non-azeotrpic refrigerant mixture.

Description

Non-azeotropic mixed refrigerant, and refrigerating apparatus using the same}

The present invention relates to a refrigeration apparatus using a non-azeotropic mixed refrigerant.

The refrigeration apparatus has a cavity for maintaining a predetermined space at a low temperature. The refrigeration apparatus is provided with a refrigeration cycle to keep the cavity at a low temperature. In the refrigeration cycle, the refrigerant circulates through the processes of compression, condensation, expansion, and evaporation.

There are various types of refrigerants, and the present disclosure targets a mixed refrigerant in which two or more types of refrigerants are mixed.

The mixed refrigerant includes an azeotropic mixed refrigerant and a non-azeotropic mixed refrigerant.

The azeotropic mixed refrigerant is a refrigerant that changes phase without changing the composition of the gas phase and the liquid phase, like a single refrigerant. In the azeotropic mixed refrigerant, the evaporation temperature is constant between the inlet of the evaporator and the outlet of the evaporator.

In the non-azeotropic mixed refrigerant, a refrigerant having a low boiling point evaporates first, and a refrigerant having a high boiling point evaporates later. Accordingly, the non-azeotropic mixed refrigerant has a different composition of the gas phase and the liquid phase, and the evaporation temperature is low at the evaporator inlet and high at the evaporator outlet.

The non-azeotropic mixed refrigerant has a Gliding Temperature Difference (GTD), which is a characteristic in which the temperature changes at isobaric pressure during a phase change.

When the non-azeotropic mixed refrigerant is used, the temperature rises when evaporation occurs at isobaric pressure, and conversely, the temperature decreases during the isostatic condensation process. In other words, a temperature gradient of the refrigerant occurs when it changes from a saturated liquid to a saturated gas state.

Using such a phenomenon, it is possible to improve the thermal efficiency of the heat exchanger. For example, the non-azeotropic mixed refrigerant may constitute a Lorentz cycle in which the temperature between the refrigerant and the heat source is balanced, and the irreversibility of heat exchange may be reduced, thereby improving efficiency.

As a conventional technology for applying the non-azeotropic mixed refrigerant, the applicant has proposed a capillary structure of a refrigeration cycle for a refrigerator KR0119839.

KR0119839 Capillary structure of refrigeration cycle for refrigerator

Although the thermal efficiency can be improved by using the non-azeotropic mixed refrigerant, an optimized composition of the non-azeotropic mixed refrigerant suitable for a refrigeration cycle applied to a refrigeration device is unknown.

A refrigeration apparatus using a non-azeotropic mixed refrigerant according to the present invention to solve the above problems includes: a compressor which is operated in a continuous operation mode to compress the non-azeotropic mixed refrigerant; A condenser for condensing the refrigerant compressed by the compressor; An expander for expanding the refrigerant condensed in the condenser; And an evaporator for evaporating the refrigerant expanded in the expander, and the pressure difference ΔP of the non-azeotropic mixed refrigerant may have a value included in the range of 340 kPa <ΔP <624.7 kPa. Accordingly, it is possible to more smoothly solve the problem of friction that occurs when the piston operates in the compressor. This operation mode can obtain a greater advantage when the compressor is operated in a continuous operation mode during operation of the refrigerating device. The continuous operation mode is an operation mode corresponding to the intermittent operation mode, and indicates a state in which the compressor is continuously operated without turning off the compressor even if the current interior temperature is provided within a target temperature range.

The condensation pressure (Pd) of the non-azeotropic refrigerant is set to have a certain value included in the range of 393.4 kPa <Pd <745.3 kPa, so that the condensation pressure realized by the non-azeotropic refrigerant can be suitably utilized in the compressor.

The evaporation pressure (Ps) of the non-azeotropic mixed refrigerant is set to have a certain value included in the range of 53.5 kPa <Ps <120.5 kPa, so that the evaporation pressure realized by the non-azeotropic mixed refrigerant can be suitably utilized in the compressor.

In the continuous operation mode, the compressor is a mode in which the internal temperature is satisfied even in a temperature range, and the operation of the piston in the compressor can be continuously performed, and in the continuous operation mode, oil circulation and gas bearings are caused by a high pressure difference. The action of the piston of the floating pressure can be obtained more faithfully.

When the compressor is a linear compressor, it is possible to reliably perform a frictional force reduction effect in driving the linear compressor by further utilizing the action of the pressure difference of the non-azeotropic mixed refrigerant.

The linear compressor includes a shell provided with a suction unit; A cylinder provided inside the shell and forming a compression space for a refrigerant; A frame coupled to the outside of the cylinder; A piston provided so as to reciprocate in the axial direction inside the cylinder; A discharge valve that is movably coupled to the cylinder and selectively discharges the refrigerant compressed in the compression space of the refrigerant; And a non-azeotropic mixed refrigerant that extends into the space between the cylinder and the frame and includes a flow path through which at least some of the refrigerants discharged from the discharge valve flow, more smoothly performs the lubricating action of the piston. can do.

In the cylinder, a cylinder body in which the nozzle part is formed; And a cylinder flange portion extending radially outward from the cylinder body, wherein the frame includes: a frame body surrounding the cylinder body; A depression into which the cylinder flange portion is inserted; And a seating portion facing the seating surface of the cylinder flange portion, so that the internal configuration of the linear compressor may be firmly supported.

The flow path may include a first flow path formed between an outer circumferential surface of the cylinder flange portion and an inner circumferential surface of the recessed portion to provide a flow path through which the compressed high-pressure non-azeotropic refrigerant is bypassed.

In the passage, a second passage formed between the seating surface of the cylinder flange portion and the seating portion of the frame may be provided, so that a high-pressure non-azeotropic mixed refrigerant passing through the first passage may be guided.

The flow path includes a third flow path extending from the second flow path to a space between the outer circumferential surface of the cylinder body and the inner circumferential surface of the frame body. Refrigerant can be guided.

Each of the flow paths may be provided with an assembly tolerance, and in this case, there is no difficulty in passing the gaseous non-azeotropic mixed refrigerant.

The linear compressor includes: a cylinder in which a cylinder step portion is formed on an inner circumferential surface; A piston formed on an outer circumferential surface of the cylinder so as to be linearly reciprocated, and forming a low pressure between the stepped portions of the cylinder when moving in one direction and forming a high pressure between the stepped portions of the cylinder when moving in the other direction; An oil suction passage formed to allow oil to flow between the cylinder stepped portion and the piston stepped portion; And an oil discharge passage formed so that the oil between the stepped portion of the cylinder and the stepped portion of the piston can be discharged to the outside of the cylinder. In the case of the linear compressor, oil circulation may be made more smoothly due to a high pressure difference between the non-azeotropic mixed refrigerant. Accordingly, the reliability of a refrigeration apparatus using a non-azeotropic mixed refrigerant can be improved.

The non-azeotropic mixed refrigerant is composed of the first hydrocarbon and the second hydrocarbon, the first hydrocarbon is isobutane, and the second hydrocarbon is provided as propane, thereby obtaining an optimum temperature gradient, thereby obtaining a high-efficiency refrigeration device. Can be obtained.

As the isobutane is provided in a weight ratio of 76% ≤ isobutane ≤ 87%, the minimum compression day of the refrigeration cycle, compatibility of refrigeration equipment production facilities, low cost of refrigerant purchase, high safety of refrigeration equipment, and increase in refrigeration cycle efficiency. , And the convenience of handling the refrigerant can be obtained.

In accordance with another aspect of the present invention, there is provided a refrigeration apparatus using a non-azeotropic mixed refrigerant comprising: a linear compressor for compressing a non-azeotropic mixed refrigerant; A condenser for condensing the non-azeotropic mixed refrigerant compressed by the compressor; An expander for expanding the non-azeotropic mixed refrigerant condensed in the condenser; And an evaporator for evaporating the non-azeotropic refrigerant expanded in the expander, and the pressure difference (ΔP) of the non-azeotropic refrigerant is 340 kPa <ΔP <624.7 kPa by using a non-azeotropic refrigerant, a linear compressor The friction reduction action between the piston and the cylinder can be reliably performed.

The linear compressor includes: a piston reciprocating; And a cylinder guiding the piston, and the high-pressure non-azeotropic refrigerant compressed by the piston is guided to the inner surface of the cylinder, so that the outer surface of the piston may float on the inner surface of the cylinder. In this case, the operating reliability of the linear compressor can be improved by the floating pressure due to the high pressure difference of the non-azeotropic mixed refrigerant.

The non-azeotropic mixed refrigerant includes at least two hydrocarbons, and the at least two hydrocarbons include at least one first hydrocarbon selected from a phase group having an evaporation temperature of -12 degrees or more at 1 bar; And at least one second hydrocarbon selected from the middle group having an evaporation temperature of -50°C or more and less than -12°C at 1 bar, and a temperature gradient difference of 4°C or more, so that a non-azeotropic mixed refrigerant is used. It can increase the cycle efficiency of and improve the operation stability of the refrigeration device.

Since the weight ratio of the first hydrocarbon is 50% or more, it is possible to optimize the compression work of the compressor.

In the linear compressor, a piston for compressing the non-azeotropic refrigerant by reciprocating motion; And a cylinder guiding the piston, and when oil pumped by a pressure difference of the non-azeotropic refrigerant is placed on a contact surface between the piston and the cylinder, abnormal supply of oil can be prevented, and Lubrication can be performed more smoothly.

In another aspect, a refrigeration apparatus using a non-azeotropic mixed refrigerant includes: a compressor controlled to compress the non-azeotropic mixed refrigerant; A condenser for condensing the refrigerant compressed by the compressor; An expander for expanding the refrigerant condensed in the condenser; And an evaporator that evaporates the refrigerant expanded in the expander to provide cool air in the chamber, wherein the compressor is operated by selecting an intermittent operation mode and a continuous operation mode, and in the continuous operation mode, the temperature in the chamber is set to a target. The compressor can be controlled to operate continuously even when within a range of temperatures. According to this, in the case of the non-azeotropic mixed refrigerant, a large pressure difference can be obtained, so that the components operated during the operation of the compressor in the continuous operation mode can operate more reliably without external influences due to frictional force.

According to the present invention, a refrigerating device capable of obtaining high efficiency can be obtained when a non-azeotropic mixed refrigerant is used.

1 is a schematic temperature graph of a non-azeotropic mixed refrigerant and air in a counterflow evaporator.
2 is a graph showing the temperature difference between the inlet and the outlet of the evaporator and the temperature gradient difference of the non-azeotropic refrigerant according to the composition of isobutane and propane.
3 is a graph showing a refrigeration cycle when (a) isobutane is used in parallel, and (b) is a non-azeotropic mixed refrigerant.
4 is a view showing a refrigerator according to the embodiment.
5 and 6 are cross-sectional views of a linear compressor applied to a refrigeration apparatus according to an embodiment, and FIG. 5 is a view showing when the piston is retracted, and FIG. 6 is a view showing when the piston is advanced.
7 is a cross-sectional view showing the configuration of an oilless linear compressor according to another embodiment.
8 is a cross-sectional view showing the configuration of a suction muffler according to the embodiment.
9 is a view showing a state in which a first filter is coupled to a suction muffler according to an embodiment.
10 is a view showing the configuration around the compression chamber according to the embodiment.
11 is an exploded perspective view showing a combination of a cylinder and a frame according to an embodiment.
12 is an exploded perspective view showing the configuration of a cylinder and a frame according to the embodiment.
13 is an exploded perspective view of a frame according to the embodiment.
14 is a cross-sectional view showing a combination of a cylinder and a piston according to an embodiment.
15 is a view showing the configuration of a cylinder according to the embodiment.
Fig. 16 is an enlarged cross-sectional view of "A" in Fig. 14;
17 is a cross-sectional view showing a combination of a frame and a cylinder according to an embodiment.
Fig. 18 is an enlarged view of "B" in Fig. 17;
19 is a cross-sectional view showing a flow of a refrigerant in a linear compressor according to an embodiment.
20 is a view showing the flow of refrigerant discharged from the compression chamber in the first and second flow paths according to the embodiment.
21 is a view showing the flow of refrigerant in a third flow path.

Hereinafter, specific embodiments of the present invention will be described in detail with reference to the drawings. However, the spirit of the present invention is not limited to the embodiments presented below, and those skilled in the art who understand the spirit of the present invention can add, change, delete, and add components to other embodiments included within the scope of the same idea. It will be possible to propose easily, but it will be said that this is also included in the idea of the present invention.

First, a non-azeotropic mixed refrigerant that can be preferably applied is presented. In the description related to the selection of the non-azeotropic mixed refrigerant, each content of the invention is divided by technical elements and described in detail. First, a process of selecting the type of non-azeotropic mixed refrigerant will be described.

<Selection of type of non-azeotropic mixed refrigerant>

It proposes mixed refrigerants suitable for the non-azeotropic mixed refrigerant. The refrigerant to be mixed was selected as a target of a hydrocarbon-based (HC-based) refrigerant. This is because the hydrocarbon-based refrigerant is an eco-friendly refrigerant having a low ozone depletion potential (ODP) and a global warming potential (GWP).

Criteria for selecting a refrigerant suitable for the non-azeotropic refrigerant among the hydrocarbon-based refrigerants can be summarized as follows.

First, from the viewpoint of the compression work, the difference (pressure difference (ΔP)) between the condensing pressure (Pd or p1) and the evaporation pressure (Ps or p2) should be small, so that the compression work of the compressor becomes small, which is advantageous for efficiency. Therefore, it is preferable to select a refrigerant having a low condensation pressure and a high evaporation pressure of the refrigerant. However, in consideration of the reliability of the compressor, it is desirable to select an evaporation pressure that satisfies 50 kPa or more.

Second, from the viewpoint of utilization of production facilities, it is preferable to use a refrigerant that has been used in the past for compatibility with conventional equipment and parts.

Third, from the viewpoint of the purchase cost of the refrigerant, it is desirable that the refrigerant can be obtained inexpensively.

Fifth, from the viewpoint of safety, a refrigerant that does not have any human obstacles when refrigerant leaks is preferable.

Sixth, from the viewpoint of reducing the irreversible loss, it is preferable to reduce the temperature difference between the refrigerant and the cold air from the viewpoint of increasing the efficiency of the cycle.

Seventh, from a handling point of view, it is desirable that the refrigerant can be conveniently handled during work, and that the operator can conveniently inject the refrigerant when the refrigerant is injected.

The selection criteria for each of the refrigerants are variously applied in selecting the refrigerant of the non-azeotropic mixed refrigerant.

<Selection of classification of hydrocarbons>

Candidate refrigerants proposed by the US National Institute of Standards and Technology were classified into three (high, medium, and low) in the order of evaporation temperature from high to low, based on evaporation temperature (Tv). As for the density of the refrigerant, the higher the evaporation temperature, the higher the density.

Among them, it is desirable to select a combination capable of producing an evaporation temperature of -20 degrees to -30 degrees suitable for the refrigerator environment. Hereinafter, the classification of the candidate refrigerants will be described in detail.

The candidate refrigerants were classified into three types based on the boundary values of evaporation temperatures of -12 degrees and -50 degrees. The candidate refrigerants classified into three types are shown in Table 1. It can be seen that the division of the evaporation temperature greatly differs based on the threshold value.

number Classification Hydrocarbon name Evaporation temperature (1 bar) Evaporation temperature (20 bar) Triple point temperature Mr. Do One Prize isopentane 27.5 154.7 -159.85 2 1,2-butadiene 10.3 124.8 -136.25 3 n-butane -0.9 114.5 -138.25 4 butene -6.6 105.8 -185.35 5 isobutane -12 100.7 -159.65 6 medium propadiene -34.7 68.2 -136.25 7 propane -42.4 57.3 -187.71 8 propylene -47.9 48.6 -185.26 9 Ha ethane -88.8 -7.2 -182.80 10 ethylene -104 -29.1 -169.15

Referring to Table 1, as the non-azeotropic refrigerant, a refrigerant may be selected and combined in each possible region as a refrigerant to be mixed. First, it looks at which classification is to be selected among the three classifications. In one case where the refrigerant mixture is selected from all three classes and three refrigerants are mixed, there may be three cases in which the refrigerant mixture is selected from two classes and two refrigerants are mixed.

When at least one refrigerant is selected from three types of refrigerants and three or more refrigerants are mixed, the rise and fall of the temperature in the non-azeotropic mixed refrigerant may become excessively large. In this case, the design of the refrigeration system is difficult, which is not desirable.

It is preferable to obtain a non-azeotropic mixed refrigerant by selecting at least one refrigerant selected from each of the two classes.

At least one refrigerant may be selected from each of the middle and lower categories, the upper and middle categories, and the upper and lower categories. Among them, it is preferable to provide a non-azeotropic mixed refrigerant as a composition in which at least one refrigerant is mixed from among the upper and middle refrigerants.

When at least one refrigerant from among the medium and sub-class refrigerants is mixed, the evaporation temperature of the refrigerant is low and too low, so that the difference between the inside temperature and the evaporation temperature of the refrigerant is too large in a general refrigerator. Accordingly, the efficiency of the refrigeration cycle deteriorates and power consumption increases, which is not preferable.

When at least one of the upper and lower refrigerants is mixed, the difference in evaporation temperature between the at least two refrigerants to be mixed is too large. Therefore, each refrigerant is classified into a liquid phase and a gas phase under actual conditions of use, unless a special high-pressure environment is created. For this reason, it is difficult to inject the at least two refrigerants into the refrigerant pipe together.

<Selection of hydrocarbons in the classification of hydrocarbons>

It will be described which refrigerant is selected from the phase classification and the intermediate classification.

First, a refrigerant selected from the phase classification will be described. At least one refrigerant selected from the phase classification may be used as the non-azeotropic mixed refrigerant.

Since the isopentane and butadiene have a relatively high evaporation temperature, there may be a problem in that the temperature inside the refrigerator evaporator is limited and the refrigeration efficiency is deteriorated.

The isobutane and N-butane can be used without changing parts of a refrigeration cycle such as a compressor of a refrigerator currently used. Therefore, it is expected to be used most among the refrigerants included in the phase classification.

The N-butane has a smaller compression work than isobutane, but has a low evaporation pressure (Ps), which may cause a problem in the reliability of the compressor. Therefore, it is not preferable compared to isobutane.

For this reason, isobutane may be most preferably selected in the phase classification. Of course, as already explained, it is not impossible to select at least one of the other hydrocarbons included in the phase classification.

A refrigerant selected from the above sub-categories will be described. At least one refrigerant selected from the medium classification may be used as the non-azeotropic mixed refrigerant.

The propadiene has an advantage of having a smaller pressure difference (ΔP) than the propane and thus has high efficiency, but is not preferable because it is expensive and has a problem that is harmful to the respiratory tract and skin when inhaled by the human body due to leakage.

The propylene is not preferable because the pressure difference is larger than that of the propane, so that the compression work of the compressor is increased.

For this reason, propane may be most preferably selected in the middle category. Of course, as already described, it is not impossible to select at least one of the other hydrocarbons included in the sub-class.

For reference, isobutane may also be referred to as R600a, and propane may also be referred to as R290. Although the isobutane and the propane have been preferably selected, it goes without saying that other hydrocarbons belonging to the same classification may be applied in obtaining the properties of the non-azeotropic mixed refrigerant. The same is true even when there is no specific mention in the following description. For example, if the temperature gradient difference of a similar non-azeotropic mixed refrigerant can be obtained, a composition other than the composition of isobutane and protan may be used.

<Selection of the ratio of the selected hydrocarbon refrigerant in consideration of the power consumption of the compression day>

The refrigerant to be mixed in the non-azeotropic mixed refrigerant was selected as isobutane in the upper category and propane in the medium category. The ratio of each refrigerant to be mixed with the non-azeotropic mixed refrigerant may be selected as follows.

The pressure difference determines the power consumption of the compressor, which is the main energy consumption source of the refrigeration system. In other words, the larger the pressure difference, the more compression work must be consumed. The larger the compression day, the worse the cycle efficiency.

The isobutane has a smaller pressure difference (ΔP) than that of the propane. For this reason, it is preferable that the weight ratio of the isobutane is 50% or more and the weight ratio of propane is 50% or less to provide the non-azeotropic mixed refrigerant.

In the case of a composition in which the non-azeotropic mixed refrigerant is mixed with isobutane and propane at 5:5, the condensation pressure is 745.3 kPa, the evaporation pressure is 120.5 kPa, and the pressure difference is 624.7 kPa.

In the case of a composition in which the non-azeotropic mixed refrigerant contains substantially all of isobutane and a very small amount of propane, the condensation pressure is 393.4 kPa, the evaporation pressure is 53.5 kPa, and the pressure difference is 340.0 Pa.

The measurement of the pressure is a measurement of an average value when the compressor is turned on under the power consumption measurement conditions of ISO. All values related to the composition of the non-azeotropic mixed refrigerant were obtained under the same conditions.

Through the above description, the range of the condensation pressure, the evaporation pressure, and the pressure difference of the non-azeotropic mixed refrigerant can be known by using the mixing ratio of isobutane and propane that can reduce the compression work.

<Selection of the selected proportion of hydrocarbon refrigerant in consideration of the irreversible loss of the evaporator>

As already described, the non-azeotropic mixed refrigerant has a temperature gradient difference (GTD) upon phase change. By using the temperature gradient difference, the evaporators are sequentially installed in each of the freezing chamber and the refrigerating chamber, thereby providing an appropriate temperature atmosphere for each partitioned space. According to the temperature gradient difference, it is possible to reduce irreversibility due to heat exchange by reducing a temperature difference between the refrigerant vaporized in each evaporator and air. It goes without saying that the loss of the refrigeration system can be reduced by reducing the irreversible loss.

1 is a schematic temperature graph of a non-azeotropic mixed refrigerant and air in a counterflow evaporator. In FIG. 1, the horizontal axis represents the traveling distance, and the air and the non-azeotropic mixed refrigerant proceed in opposite directions, respectively, as indicated by arrows. In Figure 1, the vertical axis represents the temperature.

Referring to FIG. 1, a diagram of the air (1), a diagram of the non-azeotropic refrigerant (2), a rising diagram (3) and a descending diagram of the non-azeotropic refrigerant (4), and a diagram of a single refrigerant ( 5) is shown.

Referring to the diagram of the air (1), the air may pass through the evaporator by starting in the range of -20 to -18 degrees Celsius for example, and decreasing the temperature.

Referring to the diagram (2) of the non-azeotropic mixed refrigerant, for example, the non-azeotropic mixed refrigerant may start at -27°C and increase in temperature to pass through the evaporator. The non-azeotropic mixed refrigerant may have a temperature gradient difference depending on the ratio of the isobutane and the propane. When the temperature gradient difference increases, the diagram (2) of the non-azeotropic mixed refrigerant will move toward the rising line (3) of the non-azeotropic mixed refrigerant. When the temperature gradient difference becomes small, the diagram (2) of the non-azeotropic mixed refrigerant will move toward the descending diagram (4) of the non-azeotropic mixed refrigerant.

For reference, since there is no phase change in the case of a single refrigerant, the diagram 5 of the single refrigerant may refer to that there is no temperature change.

The irreversible loss when heat exchange occurs is a phenomenon that cannot be avoided due to the temperature difference between the two interfaces for heat exchange. For example, there is no irreversible loss if there is no temperature difference at the interface between the two objects to be heat-exchanged, but heat exchange does not occur at this time.

However, there are various ways to reduce irreversible loss due to heat exchange, and a representative solution is to configure a heat exchanger with counter flow. The counterflow heat exchanger may reduce irreversible losses by reducing the temperature difference between moving fluids as much as possible.

Even in the case of an evaporator to which the non-azeotropic mixed refrigerant is applied, a heat exchanger may be configured with a counter flow as shown in FIG. 1. As the temperature of the non-azeotropic mixed refrigerant increases during evaporation due to the temperature gradient difference, the temperature difference between the air and the non-azeotropic mixed refrigerant may be reduced. When the temperature gradient difference of the non-azeotropic mixed refrigerant and the temperature difference between the air are reduced, irreversible loss may be reduced and the efficiency of a refrigeration cycle may be increased.

The temperature gradient difference of the non-azeotropic mixed refrigerant cannot be infinitely increased due to the limitation of the refrigerant. In addition, when the temperature gradient difference of the non-azeotropic mixed refrigerant is changed, the temperature gradient difference of the cold air changes, thereby changing the size of the evaporator, and further affecting the overall efficiency of the refrigeration cycle. For example, if the temperature gradient difference is increased, the inlet temperature of the refrigerant is lowered or the outlet temperature is overheated, thereby reducing the efficiency of the refrigeration cycle.

On the other hand, the temperature gradient difference of the non-azeotropic mixed refrigerant and the temperature difference between the air may converge to zero if the heat exchanger is infinitely large. However, in consideration of the mass production and economical aspects of the heat exchanger, in the case of a general refrigerator, the temperature difference between the temperature gradient of the non-azeotropic refrigerant and the air is about 3-4 degrees Celsius.

2 is a graph showing the temperature difference between the inlet and the outlet of the evaporator and the temperature gradient difference of the non-azeotropic mixed refrigerant according to the composition of the isobutane and the propane. The horizontal axis represents the content of isobutane, and the vertical axis represents the temperature difference.

Referring to FIG. 2, when 100% of the isobutane and the propane are each contained, there is no temperature change during evaporation as a single refrigerant.

When the isobutane and the propane are mixed, there is a difference in temperature gradient of the non-azeotropic mixed refrigerant and a difference in temperature between the inlet and outlet of the evaporator. The temperature difference 11 between the inlet and outlet of the evaporator is smaller than the temperature gradient difference 12 of the non-azeotropic mixed refrigerant. This may be due to incomplete heat transfer between the refrigerant and air.

The temperature gradient difference of the non-azeotropic mixed refrigerant is larger than the temperature difference at the inlet and outlet of the evaporator, from the viewpoint of making good use of the characteristics of the non-azeotropic mixed refrigerant, reducing the cost of heat exchange, and increasing the efficiency of the refrigeration cycle. desirable. Likewise, the temperature gradient difference of the non-azeotropic mixed refrigerant is preferably larger than the temperature difference of air passing through the evaporator.

In a typical refrigerator, the temperature difference of the air passing through the inlet and outlet of the evaporator may reach 4 to 10 degrees Celsius, and in most cases it is close to 4 degrees Celsius. For this reason, the temperature gradient difference of the non-azeotropic mixed refrigerant can be maintained higher than 4 degrees Celsius. More preferably, it is desirable to maintain at least 4.1 degrees Celsius, which is at least as high as the difference in temperature at the inlet and outlet of the evaporator. If the temperature gradient difference of the non-azeotropic mixed refrigerant is less than 4.1 degrees Celsius, it is not preferable because the thermal efficiency of the refrigeration cycle decreases.

On the contrary, if the temperature gradient difference of the non-azeotropic mixed refrigerant is greater than 4.1 degrees Celsius, the temperature difference between the air and the refrigerant at the outlet side of the refrigerant decreases, the irreversibility decreases, and the thermal efficiency of the refrigeration cycle increases. Here, the fact that the temperature difference between the air and the refrigerant at the outlet side of the coolant becomes small means that the diagram (2) of the non-azeotropic mixed refrigerant in FIG. 1 moves toward the rising line (3) of the non-azeotropic mixed refrigerant. I can.

When a place where the temperature gradient difference of the non-azeotropic mixed refrigerant is 4.1 degrees Celsius is found in FIG. 2, it can be seen that the isobutane is 90%, and the isobutane content is 90% or less if the temperature gradient difference is 4.1 degrees Celsius or more. Meanwhile, in order to minimize the compression work of the compressor, the isobutane is preferably 50% or more.

As a result, the weight ratio of the isobutane and the non-azeotropic mixed refrigerant provided as the propane can be proposed as shown in Equation 1.

Here, propane is the remaining weight ratio of the non-azeotropic mixed refrigerant.

As the temperature gradient difference of the non-azeotropic mixed refrigerant increases, it is preferable to reduce irreversible loss. However, when the temperature gradient difference becomes too large, the size of the evaporator becomes too large in order to secure a sufficient heat exchange path between the refrigerant and air. The evaporator applied to the general household refrigerator must be designed with a capacity of 200W or less to secure the space inside the refrigerator. For this reason, it is preferable to limit the temperature gradient difference of the non-azeotropic mixed refrigerant to 7.2 degrees Celsius or less.

In addition, if the temperature gradient difference of the non-azeotropic mixed refrigerant is too large, the temperature at the inlet of the evaporator may be too low or the evaporator outlet may be overheated too quickly based on the non-azeotropic mixed refrigerant. In this case, the usable area of the evaporator may be reduced and the heat exchange efficiency may decrease.

In detail, at the outlet of the non-azeotropic mixed refrigerant from the evaporator, the temperature of the non-azeotropic mixed refrigerant must be higher than the temperature of the air introduced into the evaporator. Otherwise, the efficiency of the heat exchanger decreases due to reversal of the temperature of the refrigerant and air. If this condition is not satisfied, the efficiency of the refrigeration system may be lowered.

When the temperature gradient difference of the non-azeotropic mixed refrigerant is found in FIG. 2, it can be seen that the isobutane is 75%, and the isobutane is 75% or more when the temperature gradient is less than 7,2 degrees Celsius.

As a result, when this condition and the condition of Equation 1 are considered together, as a result, a more preferable weight ratio of the isobutane and the non-azeotropic mixed refrigerant provided as the propane can be proposed as shown in Equation 2.

Here, propane is the remaining weight ratio of the non-azeotropic mixed refrigerant.

<Selection of the selected proportion of hydrocarbon refrigerant in consideration of compatibility of production facilities and parts>

The temperature difference between the inlet and outlet of the evaporator of a general refrigerator is set to 3-5 degrees Celsius. This is due to various factors such as parts of the refrigerator, the internal volume of the machine room, the heat capacity of each part, and the size of the fan. When the composition ratio of the non-azeotropic mixed refrigerant capable of providing the temperature of the evaporator inlet and outlet, that is, 3-5 degrees Celsius, is found in FIG. 2, it can be seen that the isobutane is between 76% and 87%.

As a result of the above discussion, it may be given by Equation 3 of the non-azeotropic mixed refrigerant that satisfies all the above-described conditions.

Here, propane is the remaining weight ratio of the non-azeotropic mixed refrigerant.

<Ratio of hydrocarbon refrigerant finally applied>

The most preferable application range of isobutane that can be selected based on the above various criteria may be determined as 81 to 82%, which is the middle range of Equation 3 above. Of course, propane may occupy other than that in the non-azeotropic mixed refrigerant.

The case of using only the isobutane and the case of using a non-azeotropic mixed refrigerant using 85% of the isobutane and 15% of the propane were compared. In both cases, the evaporator was configured in parallel to form a refrigeration system cycle.

The conditions of the experiment are the evaporator inlet temperature -29 degrees Celsius and -15 degrees Celsius, respectively, and the compressor suction temperature is 25 degrees Celsius. Due to the difference in the refrigerant, the temperature of the condenser is 31°C when only isobutane is used, and 29°C when a non-azeotropic mixed refrigerant is used.

3 is a table comparing refrigeration cycles in each case, where (a) is a case where only isobutane is used in parallel, and (b) is a case where a non-azeotropic mixed refrigerant is used.

In the experiment according to FIG. 3, it was found that there is an improvement in the coefficient of performance of approximately 4.5% when the non-azeotropic mixed refrigerant is used.

4 is a view showing a refrigeration apparatus according to an embodiment.

Referring to FIG. 4, the refrigerating apparatus according to the embodiment may include a machine room 631, a freezing room 632, and a refrigerating room 633. The refrigeration device constitutes a refrigeration cycle for operating a non-azeotropic mixed refrigerant, which has already been described. In the refrigeration cycle, a compressor 621 for compressing a refrigerant, an expander 622 for expanding the compressed refrigerant, a condenser 623 for condensing the expanded refrigerant, and an evaporator 624 and 625 for evaporating the condensed refrigerant. May be included.

The compressor 621, the expander 622, and the condenser 623 may be provided in the machine room 631. The first evaporator 624 may be provided in the freezing chamber 632. The second evaporator 625 may be provided in the refrigerating chamber 633. The freezing chamber and the refrigerating chamber may also be referred to as an interior space.

The non-azeotropic mixed refrigerant is at a lower temperature in the first evaporator 624 than in the second evaporator 625. Since the first evaporator 624 is placed in the freezing chamber, it is possible to operate the refrigeration system more suitably to the partitioned space of the refrigerating device. As a result, irreversible losses in the evaporation operation of the evaporator can be further reduced.

Hereinafter, a linear compressor that can be applied to the compressor 621 provided in the refrigerating device will be described.

5 and 6 are cross-sectional views of a linear compressor applied to a refrigeration apparatus according to an embodiment, and FIG. 5 is a time when the piston is retracted, and FIG. 6 is a time when the piston is advanced.

In the linear compressor according to the embodiment, as shown in FIGS. 5 and 6, the linear compression unit 60 is installed to be buffered inside the shell 50.

The shell 50 includes a lower shell 51 with an open top and an upper shell 52 mounted to cover the upper side of the lower shell 51. An enclosed space is formed between the lower shell 51 and the upper shell 52, and oil O is contained in the lower inner side of the lower shell 51.

The oil allows the piston to lubricate at the interface of the cylinder. If the lubrication is not smooth, it may cause a big problem in the reliability of the linear compressor.

In the shell 50, a suction pipe 53 through which fluid is sucked is disposed, and a discharge pipe 54 through which the fluid compressed by the linear compression unit 22 is discharged is disposed through.

The linear compression unit 60 is mounted on a damper 55 installed on the lower shell 51 and supported so as to be able to vibrate.

The linear compression unit 60 is arranged to be linearly reciprocated to the cylinder block 66 in which the cylinder 62 is installed, the back cover 74 provided with the fluid suction pipe 72, and the cylinder 62 A piston 80 having a fluid suction flow path 78 and a suction port 79 formed so that fluid is sucked into the cylinder 62, and a suction installed on the piston 80 to open and close the fluid suction flow path 78 A compression chamber (C) is formed between the valve 82 and the linear motor 84 for linearly reciprocating the piston 80 and the piston 80, and the fluid inside the compression chamber C is When compressed to a predetermined pressure or higher, a discharge valve assembly 92 for discharging the compressed fluid to the loop pipe 54 may be included.

The cylinder 62 is installed in the center of the cylinder block 66.

The back cover 74 is mounted on the stator cover 152 to be described later with fastening means such as fastening bolts.

A flange portion 81 is formed at a rear end of the piston 80 to receive the driving force of the linear motor 84 by being connected to the linear motor 84 by a fastening means such as a fastening bolt.

The suction valve 82 is an elastic member fastened to the front end surface of the piston 80 with a fastening bolt, and opens and closes the suction port 79 by a pressure difference between the compression chamber C and the suction port 79 do.

The linear motor 84 includes an outer core 85 installed on the cylinder block 64, a bobbin 86 installed on the outer core 85, a coil 87 wound around the bobbin 86, and , The outer core 85 and the inner core 88 installed in the cylinder block 64 to have a predetermined air gap, and the outer core 85 and the inner core to be linearly reciprocated by the electromagnetic force formed by the coil 87 A magnet 89 positioned between 88 and a magnet frame 90 to which the magnet 105 is mounted and coupled to the flange 81 of the piston 80 to transmit a linear motion force to the piston 80 ).

The discharge valve assembly 92 includes a discharge valve 93 for opening and closing the front end of the cylinder 62, and an inner side in which the discharge valve 93 is energized by a discharge spring 94 and a fluid discharge hole 95 is formed. A discharge cover 96, an outer discharge cover 97 having a flow path formed between the inner discharge cover 96, and a connection pipe mounted on the outer discharge cover 97 and connected to the roof pipe 54 ( 98).

The linear compression unit 60 is provided with an oil suction flow path 100 so that the oil contained in the shell 50 can be introduced between the cylinder 62 and the piston 80. An oil discharge passage 10 is provided so that the oil O between the cylinder 62 and the piston 80 can be discharged to the outside of the linear compression unit 60. As shown in FIG. 5 when the piston 80 is retracted, a low pressure is formed between the cylinder 62 and the piston 80, and when the piston 80 moves forward, as shown in FIG. 6, A pumping means 20 for forming a high pressure is provided between the cylinder 62 and the piston 80.

As described above, the pumping means 20 operates by a pressure difference between a low pressure and a high pressure formed between the cylinder and the piston. When the non-azeotropic mixed refrigerant is used, a pressure difference between the low pressure and the high pressure formed between the cylinder and the piston is large. Therefore, it is possible to pump more oil with a greater force compared to a refrigeration apparatus of a single refrigerant provided with equivalent cooling power.

The description of the non-azeotropic mixed refrigerant and the pumping means 20 will be described in more detail later along with the operation of the linear compressor.

The oil suction passage 10 is immersed in the oil O contained in the shell 50 and communicates with the oil pipe 11 and the oil pipe 11 mounted on the cylinder block 64 and the cylinder block An oil cover 13 having an oil passage 12 formed therebetween 64 and a cylinder formed in the cylinder block 64 so that the oil sucked through the oil passage 12 passes through the cylinder block 64 And a block suction passage 14 and a cylinder suction passage 15 formed in the cylinder 62 so that the oil sucked through the cylinder block suction passage 14 is sucked into the pumping means 30. .

The oil discharge passage 20 includes a cylinder discharge passage 21 formed in the cylinder 62 so that the oil O inside the pumping means 30 is discharged; It is configured to include a cylinder block discharge passage 22 formed in the cylinder block 64 so that the oil discharged to the cylinder discharge passage 21 passes through the cylinder block 22 and is discharged.

The pumping means (30) includes a cylinder step (31) formed on the inner circumferential surface of the cylinder (62); When the piston 80 is retracted, a low pressure is formed between the cylinder step portion 31 and a high pressure is formed between the cylinder step portion 31 when the piston 80 advances. It is configured to include a piston step 32 formed on the outer circumferential surface.

The cylinder stepped portion 31 and the piston stepped portion 32 are formed to have an inclined portion facing each other. The cylinder 62 has an inner diameter (D1) of the front of the inclined portion and an outer diameter (D2) of the rear of the inclined portion of the piston (80). ) Is formed smaller than.

The cylinder step portion 31 and the piston step portion 32 form a cylindrical space when the piston 80 is retracted.

The linear compression unit 60 opens one side of the oil suction passage 122 by a low pressure formed in the pumping means 30 when the piston 80 is retracted, and when the low pressure is released, the oil suction passage 122 ) It is configured to further include an oil intake valve 40 to seal one side.

The oil suction valve 40 is made of an elastic member fixed to the cylinder block 66 by a fastening means such as a fastening bolt, and a part of the oil suction flow path 10, in particular, of the oil path 12 is bent. Open the entrance.

The linear compressor 60 opens one side of the oil discharge passage 20 by the high pressure formed in the pumping means 30 when the piston 80 advances, and when the high pressure is released, the oil discharge passage 20 It is configured to further include an oil discharge valve 150 to seal one side.

The oil discharge valve 150 is made of an elastic member fixed to the cylinder block 66 by a fastening means such as a fastening bolt, and the oil discharge channel 20 is partially bent and, in particular, the cylinder block discharge channel 22 ) To open the exit.

Reference numeral 152 denotes a stator cover coupled to the outer core 86 by a fastening means such as a fastening bolt so as to cover the rear surface of the outer core 85.

Reference numeral 27 denotes a first spring 29 disposed between the back cover 72 and a second spring 28 disposed between the stator cover 152, and the flange 81 of the piston 80 ) Is a spring supporter that is fixed by fastening means such as fastening bolts.

Reference numeral 160 is installed at the rear end of the piston 80 to guide the fluid sucked through the suction pipe 71 of the back cover 72 to the fluid suction flow path 78 of the piston 80 and make noise. It is a muffler that reduces the weight.

The operation of the linear compressor configured as described above is as follows.

First, when a voltage is applied to the coil 87, a magnetic field is formed around the coil 87, and the magnet 89 is linearly reciprocated by interaction with the magnetic field, and the straight line of the magnet 89 The reciprocating motion is transmitted to the piston 80 through the magnet frame 90 so that the piston 80 linearly reciprocates with the cylinder 62.

The suction valve 82 and the discharge valve 93 are opened and closed by a pressure difference before and after the compression chamber C according to the linear reciprocating motion of the piston 80, and the fluid inside the shell 50 is After passing through the fluid suction pipe 72 of the cover 74, the muffler 160, the fluid suction flow path 78 and the suction port 79 of the piston 80 in order, It is sucked, compressed by the piston 80, and then passed through the discharge valve assembly 92 and the discharge pipe 54 in order to be discharged.

As described above, while the piston 80 is linearly reciprocated and the fluid inside the shell 50 is sucked/compressed/discharged, the oil O contained in the inner lower part of the shell 50 is the pumping means 30 ) It is sucked into the pumping means 30 according to an inner pressure change, lubricated/cooled between the cylinder 62 and the piston 80, and discharged to the outside of the linear compression unit 60.

The pressure change of the pumping means 30 and the corresponding oil supply process will be described in more detail as follows.

When the piston 80 is retracted as shown in FIG. 5, the piston step 32 is separated from the cylinder step 31, and the piston step 32 and the cylinder step 31 ), a low pressure is formed, and the oil suction valve 40 is partially bent by the low pressure to open the inlet of the oil suction passage 10, in particular, the oil passage 12, and the oil discharge valve 150 seals the oil discharge passage 20, in particular, the cylinder block discharge passage 150 by the low pressure.

Oil (O) contained in the inner lower part of the shell 50 passes through the oil pipe 11, the oil passage 12, the cylinder block suction passage 14, and the cylinder suction passage 15 in order by the low pressure. Then, it is sucked into the space between the piston step 32 and the cylinder step 31 and lubricates/cools the cylinder 62 and the piston 80.

On the other hand, when the piston 80 is advanced, as shown in FIG. 6, the piston step 32 becomes closer to the cylinder step 31, and the piston step 32 and the cylinder step A high pressure is formed between the parts 31, and the oil suction valve 40 seals the inlet of the oil suction passage 10, in particular, the oil passage 12 by the high pressure, and the oil discharge valve ( 150) is partially bent by the high pressure to open the oil discharge passage 20, in particular, the cylinder block discharge passage 150.

The oil in the space between the piston stepped portion 32 and the cylinder stepped portion 31 passes through the cylinder discharge passage 21 and the cylinder block discharge passage 22 by the high pressure in order to the linear compression portion. It is discharged to the outside of 60.

As the pressure difference of the linear compressor increases, the pumping means 20 may supply a large amount of oil with a greater force. The pressure difference of the linear compressor may correspond to a pressure difference (ΔP) between the evaporation pressure and the condensation pressure of the refrigerant.

When a non-azeotropic mixed refrigerant is used as the working fluid of the linear compressor, it can operate with a larger pressure difference (ΔP) compared to the case of using isobutane, which is widely used as a single refrigerant.

The pressure difference of the non-azeotropic mixed refrigerant applied to the examples will be described. First, in the case of a composition in which isobutane and propane are mixed at 5:5, the condensation pressure is 745.3 kPa, the evaporation pressure is 20.5 kPa, and the pressure difference is 624.7 kPa. In the case of a composition in which the non-azeotropic mixed refrigerant contains substantially all of isobutane and a very small amount of propane, the condensation pressure is 393.4 kPa, the evaporation pressure is 53.5 kPa, and the pressure difference is 340.0 Pa.

Consequently, the pressure difference of the non-azeotropic mixed refrigerant of the embodiment may have a range of 340.0 kPa or more and 624.7 kPa or less.

As described above, when the non-azeotropic mixed refrigerant is used as the working fluid, oil can be supplied to the interior of the linear compressor as a larger pressure difference. In this case, it is possible to prevent a phenomenon in which the oil supply flow path is blocked, the oil circulation is delayed, or the contact surface between the piston and the cylinder is overheated.

When the non-azeotropic mixed refrigerant is used in a linear compressor, a greater advantage can be expected when the refrigeration system is operated in a continuous operation mode.

First, the continuous operation mode will be described in detail.

The refrigeration device is driven differently in a refrigeration cycle in response to a temperature range divided by an upper limit reference value and a lower limit reference value based on a set temperature in the container.

Specifically, the temperature range is divided into three as follows. First, when the inside temperature is in a satisfactory temperature range within the upper and lower reference values, second, when the inside temperature is in an unsatisfactory temperature range exceeding the upper limit reference value, and third, a subcooled temperature in which the inside temperature is less than the lower limit reference value. It can be divided into areas.

The control unit of the refrigeration unit controls the refrigeration unit to operate to supply cooling power when the interior temperature reaches the unsatisfactory temperature range, and when the interior temperature reaches the subcooling temperature range, the cooler stops supplying cooling power. Control to do it. Such an operation mode can be referred to as an intermittent operation mode.

When the upper limit reference value and the lower limit reference value are set to be narrow so that the amount of temperature change in the container is reduced, the freshness of the article is improved and the storage period of the article can be improved. This can be said to have improved the constant temperature function, and it can be said that the refrigeration device operates more preferably.

In the intermittent operation mode, in order to improve the constant temperature function, the number of times the refrigeration cycle is driven and stopped is increased. Due to this, the reliability of parts is deteriorated due to frequent on/off of refrigeration cycle parts, power consumption increases every time the parts are turned on after being turned off, and excessive cooling power is initially supplied when turning on after being turned on, resulting in a constant temperature function. This aggravates the downside.

In order to solve this problem, a continuous operation mode may be applied to reduce the number of times the cooling cycle is driven and stopped.

In the intermittent operation mode, when the temperature in the chamber reaches the upper limit reference value and the lower limit reference value, the control unit of the refrigeration self-government controls switching to start or stop supplying cooling power to the refrigerating device. When the inside temperature is in the satisfactory temperature range, the control unit controls to stop supply of cooling power to the refrigerating device.

In contrast, in the continuous operation mode, the supply of cooling power to the cooler is not stopped even when the inside of the container is in a satisfactory temperature range.

The controller may control the refrigeration device based on the current temperature and the target temperature measured by the temperature sensor. For example, the controller may adjust the cooling power of the cooler based on a difference value between the target temperature in the container and the current temperature in the container sensed by the temperature sensor. As another example, the controller may adjust the cooling power of the cooler based on an increase or decrease in the internal temperature detected by the temperature sensor at a predetermined time interval.

The control unit may adjust the cooling power of the cooler based on the cooling power of the refrigerating device previously determined. For example, the control unit may determine the cooling power to be output at the current point in time as an output of the refrigerating device as a value that fluctuates in proportion to the sum of the previously determined first and second outputs of the cooling device.

As the refrigeration device is operated in the continuous operation mode as described above, the temperature inside the container can be maintained near the target temperature without deviating to the unsatisfactory temperature region.

In the continuous operation mode, the product is performed in a frozen or refrigerated state, and in a special operation operated according to a special situation, the intermittent operation mode may be operated. For example, when the start condition of the special operation in the refrigeration unit is satisfied, the control unit of the refrigeration unit switches to the intermittent operation mode to drive the refrigerator, and when the end condition of the special operation is satisfied, the control unit enters the continuous operation mode. It can be switched and controlled to drive the refrigerator.

The special operation may include a defrost operation, a door load response operation, an operation when power is initially applied, and the like.

The defrosting operation will be described as an example.

When at least one of the case where the defrost cycle in the refrigerator compartment has elapsed and the case where the evaporator temperature in the compartment reaches a predetermined specific value is satisfied, the control unit may terminate the continuous operation mode and switch to the intermittent operation mode. have.

The operation of the refrigerating device in the defrosting operation will be described in more detail.

First, in order to pre-cool the internal temperature of the chamber that will rise during the defrosting step, the control unit performs a deep cooling step of controlling the cooler to supply cooling power greater than that in the previous continuous operation mode. I can. When a predetermined time elapses after the deep cooling step or reaches a set temperature, the control unit may end the deep cooling step.

When the deep cooling step is finished, the control unit may perform a defrost step of supplying heat to the refrigerating device by stopping the supply of cooling power to the refrigerating device and turning on a defrost heater to melt the ice deposited in the evaporator. .

When the defrosting step is finished, the temperature in the bin, which has risen during the defrosting step, must be rapidly lowered. To this end, the control unit causes the evaporator to receive a cooling power greater than that received in the continuous operation mode immediately before. This can be referred to as driving after defrost.

A point in time when the operation after the defrosting is terminated is a termination condition for the defrosting operation. When or after the defrosting operation is terminated, the control unit may control to terminate the intermittent operation mode of the refrigerator and switch to the continuous operation mode.

The door load response will be described as an example.

When at least one of the cases in which a certain period of time has elapsed after the door of the refrigerator is opened and closed or the temperature in the container has reached a predetermined specific value after the door of the refrigerator is opened and closed is satisfied, the starting condition for the door load response operation Is satisfied. At this time, the control unit controls the refrigerator to end the continuous operation mode and switch to the intermittent operation mode.

The operation of the refrigerating device in the door load response operation will be described in detail.

First, in order to remove the heat load that has flowed into the interior by opening the door, the control unit controls to perform a door load response operation in which the refrigerating device supplies cooling power greater than the cooling power in the immediately preceding continuous operation mode. I can.

When a certain period of time elapses after the door load response operation is performed, or when the temperature in the cabinet reaches a set temperature, the control unit ends the door load response operation.

When the door load response operation is terminated or after the operation is terminated, the control unit may control to terminate the intermittent operation mode of the refrigerating device and switch to the continuous operation mode.

The operation at the time of initial application of the power will be described as an example.

When power is applied to the refrigerator again after the refrigerating device is turned off, the operation start condition of the initial power supply operation is satisfied. In this case, the controller may control the refrigerator in the intermittent operation mode in preference to the continuous operation mode.

Specifically, in order to quickly lower the temperature in the chamber that has risen while the power is cut off, the control unit performs an initial power supply operation in which the evaporator supplies cooling power greater than that in the previous continuous operation mode.

When a certain period of time elapses after the initial power supply operation is performed, or when the temperature in the chamber reaches a set temperature, the control unit terminates the operation when the power is initially applied.

When the power is initially applied, the control unit may control to end the intermittent operation mode and switch to the continuous operation mode when the operation is terminated or after the operation is terminated.

When operating in the continuous operation mode, the control unit of the refrigerating apparatus may control the evaporator to receive a cooling power lower than that supplied in the intermittent operation mode. In particular, in the continuous operation mode that is performed for a longer time than the intermittent operation mode, there is a characteristic that low cooling power must be continuously supplied while being variably controlled.

In a refrigeration system that variably controls low cooling power as described above, it may be advantageous to have a low evaporation temperature of the refrigerant.

The relationship between the refrigeration system and the refrigerant evaporation temperature will be described.

When the control unit of the refrigerating device controls the cooling power of the evaporator to decrease, it means to reduce the output of the compressor to reduce the flow rate or flow rate of the refrigerant circulating inside the refrigerant pipe.

In a situation in which the refrigerant continuously circulates with low cooling power to perform the continuous operation mode, the circulation rate of the refrigerant in a refrigeration system having a relatively high evaporation temperature of the refrigerant is The flow rate or flow rate must be large. This is because if the same amount of refrigerant evaporates at a lower temperature, it can produce more cooling power.

Consequently, the compressor output in the refrigeration system having a high evaporation temperature must be greater than the compressor output in the refrigeration system having a low evaporation temperature, and likewise, there is a disadvantage in that power consumption for driving the compressor is increased. In addition, as the period in which the continuous operation mode is performed is longer, a wear resistance problem of components such as a compressor may occur.

Due to the above advantages, it is advantageous to use the non-azeotropic refrigerant mixed with a refrigerant such as propane having a relatively lower evaporation temperature than isobutane having a relatively high evaporation temperature, in the refrigeration apparatus to which the continuous operation mode is applied. Do.

In the refrigeration apparatus to which the continuous operation mode is applied, a reliability problem may occur when a compressor configured to float a piston of a compressor in a cylinder by oil is used.

As seen in the above linear compression, in the linear compressor lubricated with oil, the oil inside the compressor is circulated by the pressure difference ΔP of the refrigerant generated during the reciprocating movement of the piston. Therefore, when driving with low cooling power, at least one of the frequency and stroke of the linear compressor is lowered, so that the circulation speed of the oil and the amount of circulation of the oil decrease, and thus the oil supply flow path may be blocked.

In the continuous operation mode, since there are much more sections in which low cooling power is supplied than in sections in which high cooling power is supplied, a problem may occur in circulation of the oil. In order to solve this problem, when a non-azeotropic mixed refrigerant having a larger pressure difference (ΔP) compared to a single refrigerant exemplified by isobutane is used, it is possible to reduce a decrease in the circulation rate of oil and reduce clogging of the oil supply flow path. There will be.

It has been described that the non-azeotropic mixed refrigerant is more preferably applied to the circulation of oil for lubrication of the linear compressor.

Not only this, but the non-azeotropic mixed refrigerant can be preferably applied even in the case of a linear compressor in which oil is not used for lubrication. In this case, the linear compressor may be lubricated by air. More precisely, the contact of the piston and the cylinder can be prevented by air. Such a compressor can also be referred to as an oilless linear compressor.

The oilless linear compressor may form an air layer between the cylinder and the piston in place of oil required to prevent wear and damage between the cylinder and the piston.

The oilless compressor may cause a piston to float in a cylinder due to a pressure difference ΔP of a refrigerant generated during a reciprocating motion of the piston. Therefore, it is advantageous to apply a non-azeotropic mixed refrigerant having a large pressure difference (ΔP).

7 is a cross-sectional view showing the configuration of an oilless linear compressor according to another embodiment.

Referring to FIG. 7, the linear compressor 100 according to the embodiment includes a shell 101 having a substantially cylindrical shape, a first cover 102 coupled to one side of the shell 101, and a second cover 102 coupled to the other side. A cover 103 is included. For example, the linear compressor 100 is laid in a horizontal direction, the first cover 102 is on the right side of the shell 101, and the second cover 103 is on the left side of the shell 101. Can be combined.

Here, the linear compressor 100 exemplifies a linear compressor in which oil is not used.

In a broader sense, the first cover 102 and the second cover 103 can be understood as one configuration of the shell 101.

The linear compressor 100 is provided with a driving force to the cylinder 120 provided inside the shell 101, the piston 130 and the piston 130 reciprocating and linearly moving inside the cylinder 120 A motor assembly 140 is included as a linear motor.

When the motor assembly 140 is driven, the piston 130 may reciprocate at high speed. The operating frequency of the linear compressor 100 according to the present embodiment is approximately 100 Hz.

In detail, the linear compressor 100 includes a suction unit 104 through which refrigerant is introduced and a discharge unit 105 through which the refrigerant compressed in the cylinder 120 is discharged. The suction unit 104 may be coupled to the first cover 102, and the discharge unit 105 may be coupled to the second cover 103.

The refrigerant sucked through the suction part 104 flows into the piston 130 through the suction muffler 150. In the process of passing the refrigerant through the suction muffler 150, noise may be reduced. The suction muffler 150 is configured by combining a first muffler 151 and a second muffler 153. At least a portion of the suction muffler 150 is located inside the piston 130.

The piston 130 includes a substantially cylindrical piston body 131 and a piston flange portion 132 extending radially from the piston body 131. The piston body 131 reciprocates within the cylinder 120, and the piston flange portion 132 may reciprocate outside the cylinder 120.

The piston 130 may be made of a non-magnetic aluminum material (aluminum or aluminum alloy). Since the piston 130 is made of an aluminum material, the magnetic flux generated by the motor assembly 140 is transmitted to the piston 130 to prevent leakage to the outside of the piston 130. In addition, the piston 130 may be formed by a forging method.

Meanwhile, the cylinder 120 may be made of a non-magnetic aluminum material (aluminum or aluminum alloy). In addition, the material composition ratio of the cylinder 120 and the piston 130, that is, the type and composition ratio may be the same.

Since the cylinder 120 is made of an aluminum material, the magnetic flux generated by the motor assembly 200 is transmitted to the cylinder 120 to prevent leakage to the outside of the cylinder 120. In addition, the cylinder 120 may be formed by an extrusion rod processing method.

In addition, since the piston 130 and the cylinder 120 are made of the same material (aluminum), the coefficients of thermal expansion are the same. During the operation of the linear compressor 100, a high temperature (about 100°C) environment is created inside the shell 100, and the piston 130 and the cylinder 120 have the same coefficient of thermal expansion, so that the piston 130 And the cylinder 120 can be thermally deformed by the same amount.

As a result, since the piston 130 and the cylinder 120 are thermally deformed in different sizes or directions, it is possible to prevent interference with the cylinder 120 between movements of the piston and 130.

The cylinder 120 is configured to receive at least a portion of the suction muffler 150 and at least a portion of the piston 130.

Inside the cylinder 120, a compression space P in which the refrigerant is compressed by the piston 130 is formed. In addition, a suction hole 133 for introducing a refrigerant into the compression space P is formed in a front portion of the piston 130, and the suction hole 133 is selectively disposed in front of the suction hole 133. An opening intake valve 135 is provided. A fastening hole through which a predetermined fastening member is coupled is formed in an approximately central portion of the suction valve 135.

In front of the compression space P, a discharge cover 160 forming a discharge space or discharge flow path of the refrigerant discharged from the compression space P and the discharge cover 160 is coupled to the compressed space P Discharge valve assemblies 161, 162, and 163 for selectively discharging the refrigerant compressed in FIG.

In the discharge valve assemblies (161, 162, 163), a discharge valve (161) that is opened when the pressure in the compression space (P) becomes equal to or higher than the discharge pressure to flow the refrigerant into the discharge space of the discharge cover (160), and the discharge valve ( A valve spring 162 provided between the 161 and the discharge cover 160 to impart an elastic force in the axial direction, and a stopper 163 for limiting the amount of deformation of the valve spring 162 are included. Here, the compression space P is understood as a space formed between the intake valve 135 and the discharge valve 161.

In addition, the "axial direction" may be understood as a direction in which the piston 130 reciprocates, that is, a transverse direction in FIG. 7. In the "axial direction", a direction from the suction part 104 toward the discharge part 105, that is, a direction in which the refrigerant flows, is defined as "front", and the opposite direction is defined as "rear".

On the other hand, the "radial direction" is a direction perpendicular to the direction in which the piston 130 reciprocates, and can be understood as the vertical direction of FIG. 3.

The stopper 163 may be seated on the discharge cover 160, and the valve spring 162 may be seated at the rear of the stopper 163. In addition, the discharge valve 161 is coupled to the valve spring 162, and the rear or rear portion of the discharge valve 161 is positioned to be supported on the front surface of the cylinder 120.

The valve spring 162 may include, for example, a plate spring.

The suction valve 135 may be formed on one side of the compression space P, and the discharge valve 161 may be provided on the other side of the compression space P, that is, on the opposite side of the suction valve 135.

In the course of the piston 130 reciprocating and linear movement inside the cylinder 120, when the pressure in the compression space P is lower than the discharge pressure and less than the suction pressure, the suction valve 135 is opened and the refrigerant Is sucked into the compression space (P). On the other hand, when the pressure in the compression space P is equal to or higher than the suction pressure, the refrigerant in the compression space P is compressed while the suction valve 135 is closed.

On the other hand, when the pressure in the compression space (P) is greater than or equal to the discharge pressure, the valve spring 162 is deformed to open the discharge valve 161, and the refrigerant is discharged from the compression space (P) to be discharged. It is discharged to the discharge space of the cover 160.

In addition, the refrigerant flowing through the discharge space of the discharge cover 160 flows into the roof pipe 165. The roof pipe 165 is coupled to the discharge cover 160 and extends to the discharge part 105, and guides the compressed refrigerant in the discharge space to the discharge part 105. For example, the roof pipe 178 has a shape wound in a predetermined direction and extends roundly, and is coupled to the discharge part 105.

The linear compressor 100 further includes a frame 110. The frame 110 is configured to fix the cylinder 120 and may be fastened to the cylinder 200 by a separate fastening member. The frame 110 is disposed to surround the cylinder 120. That is, the cylinder 120 may be positioned to be received inside the frame 110. In addition, the discharge cover 172 may be coupled to the front surface of the frame 110.

On the other hand, at least some of the gas refrigerant of the high-pressure gas refrigerant discharged through the open discharge valve 161 is directed toward the outer circumferential surface of the cylinder 120 through the space where the cylinder 120 and the frame 110 are combined. It can be fluid.

In addition, the refrigerant is introduced into the cylinder 120 through a gas inlet portion 122 (see FIG. 16) and a nozzle portion 123 (see FIG. 16) formed in the cylinder 120. The introduced refrigerant flows into the space between the piston 130 and the cylinder 120 so that the outer circumferential surface of the piston 130 is spaced apart from the inner circumferential surface of the cylinder 120. Accordingly, the introduced refrigerant may function as a “gas bearing” that reduces friction with the cylinder 120 during the reciprocating movement of the piston 130.

In the motor assembly 140, outer stators 141, 143, and 145 are fixed to the frame 110 and disposed to surround the cylinder 120, and an inner stator 148 that is spaced apart from the inner stator 141, 143, and 145. ) And a permanent magnet 146 positioned in the space between the outer stator 141, 143, and 145 and the inner stator 148.

The permanent magnet 146 may linearly reciprocate by mutual electromagnetic force between the outer stator 141, 143, and 145 and the inner stator 148. In addition, the permanent magnet 146 may be composed of a single magnet having one pole, or may be configured by combining a plurality of magnets having three poles.

The permanent magnet 146 may be coupled to the piston 130 by a connection member 138. In detail, the connection member 138 may be coupled to the piston flange portion 132 to extend by bending toward the permanent magnet 146. As the permanent magnet 146 reciprocates, the piston 130 may reciprocate together with the permanent magnet 146 in the axial direction.

Further, the motor assembly 140 further includes a fixing member 147 for fixing the permanent magnet 146 to the connection member 138. The fixing member 147 may be formed by mixing glass fiber or carbon fiber and resin. The fixing member 147 is provided to surround the inner and outer sides of the permanent magnet 146, so that the permanent magnet 146 and the connection member 138 may be firmly maintained in a coupled state.

The outer stator (141,143,145) includes coil winding bodies (143,145) and a stator core (141).

The coil winding bodies 143 and 145 include a bobbin 143 and a coil 145 wound in the circumferential direction of the bobbin 143. The cross section of the coil 145 may have a polygonal shape, and for example, may have a hexagonal shape.

The stator core 141 is configured by stacking a plurality of laminations in a circumferential direction, and may be disposed to surround the coil winding bodies 143 and 145.

A stator cover 149 is provided on one side of the outer stator 141, 143, and 145. One side of the outer stator 141, 143, and 145 may be supported by the frame 110, and the other side may be supported by the stator cover 149.

The inner stator 148 is fixed to the outer periphery of the frame 110. In addition, the inner stator 148 is configured by stacking a plurality of laminations from the outside of the cylinder 120 in the circumferential direction.

The linear compressor 100 further includes a supporter 137 supporting the piston 130 and a back cover 170 spring-coupled to the supporter 137.

The supporter 137 is coupled to the piston flange portion 132 and the connection member 138 by a predetermined fastening member.

In front of the back cover 170, a suction guide part 155 is coupled. The suction guide part 155 guides the refrigerant sucked through the suction part 104 to be introduced into the suction muffler 150.

The linear compressor 100 includes a plurality of springs 176 whose natural frequencies are adjusted so that the piston 130 can perform resonant motion.

The plurality of springs 176 include a first spring supported between the supporter 137 and the stator cover 149 and a second spring supported between the supporter 137 and the back cover 170 do.

The linear compressor 100 further includes leaf springs 172 and 174 provided on both sides of the shell 101 so that the internal components of the compressor 100 are supported by the shell 101.

The leaf springs 172 and 174 include a first leaf spring 172 coupled to the first cover 102 and a second leaf spring 174 coupled to the second cover 103. As an example, the first plate spring 172 may be fitted into a portion where the shell 101 and the first cover 102 are coupled, and the second plate spring 174 is 2 The cover 103 may be arranged to be fitted to a portion to which it is coupled.

8 is a cross-sectional view showing a configuration of a suction muffler according to an embodiment, and FIG. 9 is a view showing a state in which a first filter is coupled to the suction muffler according to the embodiment.

8 and 9, the suction muffler 150 includes a first muffler 151, a second muffler 153 coupled to the first muffler 151, and the first muffler 151. A first filter 310 supported by the second muffler 153 is included.

The first muffler 151 and the second muffler 153 have a flow space portion through which a refrigerant flows. In detail, the first muffler 151 extends from the inside of the suction part 104 toward the discharge part 105, and at least a portion of the first muffler 151 is inside the suction guide part 155 Is extended to. In addition, the second muffler 153 extends from the first muffler 151 to the inside of the piston body 131.

The first filter 310 is understood as a configuration installed in the flow space to filter foreign matter. Since the first filter 310 is made of a material having magnetic properties, it is possible to easily filter foreign matter contained in the refrigerant, particularly metal dirt.

As an example, the first filter 310 may be made of stainless steel, and thus may have a predetermined magnetism and a rust phenomenon may occur.

As another example, the first filter 310 may be coated with a magnetic material or configured to attach a magnet to the surface of the first filter 310.

The first filter 310 may be configured in a mesh type having a plurality of filter holes, and may have a substantially disk shape. In addition, the filter hole may have a diameter or width less than or equal to a predetermined size. For example, the predetermined size may be about 25 μm.

The first muffler 151 and the second muffler 153 may be assembled by a press-fitting method. In addition, the first filter 310 may be assembled by being fitted into a press-fit portion of the first muffler 151 and the second muffler 153.

In detail, the first muffler 151 has a groove portion 151a to which at least a portion of the second muffler 153 is coupled. Further, the second muffler 153 includes a protrusion 153a inserted into the groove 151a of the first muffler 151.

When both side portions of the first filter 310 are interposed between the groove portion 151a and the protrusion portion 153a, the first filter 310 may be supported by the first and second mufflers 151 and 153. have.

When the first filter 310 is positioned between the first and second mufflers 151 and 153, when the first muffler 151 and the second muffler 153 move in a direction closer to each other and press-fit, Both side portions of the first filter 310 may be inserted and fixed between the groove portion 151a and the protrusion portion 153a.

In this way, since the first filter 310 is provided to the suction muffler 150, foreign substances having a predetermined size or more among the refrigerant sucked through the suction unit 104 may be filtered by the first filter 310. . Therefore, foreign matters are contained in the refrigerant acting as a gas bearing between the piston 130 and the cylinder 120, and it is possible to prevent the foreign matter from flowing into the cylinder 120.

In addition, since the first filter 310 is firmly fixed to the press-fit portion of the first and second mufflers 151 and 153, separation from the suction muffler 150 can be prevented.

In the present embodiment, it has been described that the groove portion 151a is formed in the first muffler 151 and the protrusion portion 153a is formed in the second muffler 153, but unlike this, the protrusion portion 151a is formed in the first muffler 151. May be formed and a groove portion may be formed in the second muffler 153.

10 is a view showing a configuration around a compression chamber according to an embodiment, FIG. 11 is an exploded perspective view showing a combination of a cylinder and a frame according to the embodiment, and FIG. 12 is a view showing the configuration of a cylinder and a frame according to the embodiment An exploded perspective view, FIG. 13 is an exploded perspective view of a frame according to an embodiment, and FIG. 14 is a cross-sectional view showing a combination of a cylinder and a piston according to the embodiment.

10 to 14, in the linear compressor 100, at least some of the refrigerants compressed and discharged in the compression chamber P flow into the space between the frame 110 and the cylinder 120. The space between the frame 110 and the cylinder 120 is formed by the assembly tolerance of the frame 110 and the cylinder 120, between the inner surface of the frame 110 and the outer surface of the cylinder 120 Is understood as the gap of.

In the space between the frame 110 and the cylinder 120, flow paths 410, 420, and 430 are included. The passages 410, 420, and 430 include a first passage 410, a second passage 420, and a third passage 430 sequentially formed in a direction in which the refrigerant flows.

In detail, the cylinder 120 includes a cylinder body 121 having a substantially cylindrical shape and a cylinder flange portion 125 extending radially from the cylinder body 121.

The cylinder body 121 includes a gas inlet 122 through which the discharged gas refrigerant is introduced. The gas inlet 122 may be formed in a circular shape along the outer circumferential surface of the cylinder body 121.

In addition, a plurality of gas inlets 122 may be provided. The plurality of gas inlet portions 122 include gas inlet portions 122a, 122b (see FIG. 15) located on one side from the central portion in the axial direction of the cylinder body 121 and a gas inlet portion positioned on the other side from the central portion in the axial direction. (122c, see Fig. 15) is included.

The cylinder flange portion 125 is provided with a fastening portion 126 coupled to the frame 110. The fastening part 126 may be configured to protrude outward from the outer circumferential surface of the cylinder flange part 125. The fastening part 126 may be coupled to the cylinder fastening hole 118 of the frame 110 by a predetermined fastening member, for example, a bolt.

The cylinder flange portion 125 includes a seating surface 127 that is seated on the frame 110. The seating surface 127 may be a rear portion of the cylinder flange portion 125 extending in the radial direction from the cylinder body 121.

In the frame 110, a frame body 111 surrounding the cylinder body 121 and a cover coupling portion 115 extending in a radial direction of the frame body 111 and coupled to the discharge cover 160 Is included.

In the cover coupling part 115, a plurality of cover fastening holes 116 into which a fastening member coupled to the discharge cover 160 is inserted, and a plurality of cylinders into which a fastening member coupled to the cylinder flange 125 is inserted. Fastening holes 118 are formed. The cylinder fastening hole 118 is formed at a slightly recessed position from the cover coupling part 115.

The frame 110 is provided with a depression 117 communicating with the frame body 111. The recessed part 117 is formed by being recessed rearward from the cover coupling part 115, and the cylinder flange part 125 may be inserted into the recessed part 117. That is, the depression 117 may be disposed to surround the outer circumferential surface of the cylinder flange 125. The depressed depth of the depressed portion 117 may correspond to a front and rear width of the cylinder flange portion 125.

A predetermined refrigerant flow space, that is, the first flow path 410 may be formed between the inner circumferential surface of the depression 117 and the outer circumferential surface of the cylinder flange 125. When the cylinder 120 is assembled to the frame 110, a predetermined assembly tolerance is formed between the outer circumferential surface of the cylinder flange 125 and the inner circumferential surface of the depression 117, A corresponding space forms the first flow path 410.

The high-pressure gas refrigerant discharged from the discharge valve 161 passes through the first flow path 410 and flows into the second flow path 420 provided with the second filter 320. The second filter 320 may be understood to be a filter member provided between the frame 110 and the cylinder 120 to filter the high-pressure gas refrigerant discharged through the discharge valve 161.

In detail, at the rear end of the recessed portion 117, a seating portion 113 provided stepwisely is formed. The seating portion 113 extends radially inward from the recessed portion 117 and is positioned to face the seating surface 127 of the cylinder flange portion 125.

A ring-shaped second filter 320 may be mounted on the mounting portion 113.

In a state in which the second filter 320 is seated on the seating portion 113, when the cylinder 120 is coupled to the frame 110, the cylinder flange portion 125 is the second filter 320 The second filter 320 is pressed in front of. That is, the second filter 320 may be interposed and fixed between the seating portion 113 of the frame 110 and the seating surface 127 of the cylinder flange portion 125.

The second flow path 420 is a flow path through which the refrigerant passed through the first flow path 410 flows, and a predetermined value is provided between the seating portion 113 and the seating surface 127 of the cylinder flange portion 125. An assembly tolerance is formed, and a space corresponding to the assembly tolerance forms the second flow path 420.

The second filter 320 is installed in the second flow path 420 so that foreign matters from the high-pressure gas refrigerant flowing through the second flow path 420 flow into the gas inlet 122 of the cylinder 120 It is blocked from being, and may be configured to adsorb oil contained in the refrigerant.

For example, the second filter 320 may include a nonwoven fabric or an adsorption fabric made of polyethylene terephthalate (PET) fibers. The PET has the advantage of excellent heat resistance and mechanical strength. In addition, it is possible to block foreign substances of 2 μm or more in the refrigerant.

We propose another embodiment.

In the above embodiment, it has been described that the second filter 320 is installed in the second flow path 420, but unlike this, the second filter 320 is the first flow path 410, that is, the cylinder plan. It may be installed in the space between the outer circumferential surface of the branch part 125 and the inner circumferential surface of the recessed part 117 of the frame 110.

The flow paths 410, 420, and 430 include a third flow path 430 through which the refrigerant passed through the second flow path 420 flows.

The third flow path 430 extends rearward along the outer circumferential surface of the cylinder body 121 from the second flow path 420, and the rear portion of the frame body 111 and the first of the cylinder body 121 It may extend to a space between the body end portions 121a (refer to FIG. 15).

The refrigerant flowing through the third flow path 430 may flow toward the inner circumferential surface of the cylinder 120 through the gas inlet 122 and the nozzle 123.

15 is a view showing the configuration of a cylinder according to the embodiment, and FIG. 16 is an enlarged cross-sectional view of “A” in FIG. 14.

15 and 16, in the cylinder 120 according to the embodiment, a cylinder body 121 having a substantially cylindrical shape and forming a first body end 121a and a second body end 121b, and the cylinder A cylinder flange portion 125 extending radially outward from the second body end 121b of the body 121 is included.

The first body end 121a and the second body end 121b form both ends of the cylinder body 121 with respect to the central portion 121c in the axial direction of the cylinder body 121.

The cylinder body 121 includes a plurality of gas inlets 122 in which at least some of the high-pressure gas refrigerants discharged through the discharge valve 161 flow and a third filter 330 is installed. . Further, the cylinder body 121 further includes a nozzle portion 123 extending radially inward from the plurality of gas inlet portions 122.

The plurality of gas inlet portions 122 and nozzle portions 123 are understood as one configuration of the third flow path 430. Accordingly, at least some of the refrigerants flowing through the third flow path 430 may flow toward the inner circumferential surface of the cylinder 120 through the plurality of gas inlet portions 122 and nozzle portions 123.

The plurality of gas inlets 122 are configured to be depressed by a predetermined depth and width from the outer circumferential surface of the cylinder body 121.

In addition, the introduced refrigerant is located between the outer circumferential surface of the piston 130 and the inner circumferential surface of the cylinder 120, and functions as a gas bearing for the movement of the piston 130. That is, due to the pressure of the refrigerant, the outer circumferential surface of the piston 130 is kept spaced apart from the inner circumferential surface of the cylinder 120.

The plurality of gas inlets 122 include a first gas inlet 122a and a second gas inlet 122b positioned at one side from the axial center 121c of the cylinder body 121, and the shaft A third gas inlet 122c positioned on the other side from the direction central portion 121c is included.

The first and second gas inlets 122a and 122b are located closer to the second body end 121b with respect to the axial central portion 121c of the cylinder body 121, and the third gas inlet portion 122c may be positioned closer to the first body end 121a with respect to the central portion 121c in the axial direction of the cylinder body 121.

That is, the plurality of gas inlets 122 are arranged in an asymmetrical number with respect to the central portion 121c in the axial direction of the cylinder body 121.

Referring to FIG. 15, the internal pressure of the cylinder 120 is more on the side of the second body end 121b close to the discharge side of the compressed refrigerant compared to the first body end 121a close to the suction side of the refrigerant. Since it is formed high, more gas inlet portions 122 are formed on the side of the second main body end 121b to enhance the function of the gas bearing, whereas a relatively small amount of gas inlet portions on the side of the first main body end 121a ( 122).

The cylinder body 121 further includes a nozzle part 123 extending from the plurality of gas inlet parts 122 in a direction of the inner circumferential surface of the cylinder body 121. The nozzle part 123 is formed to have a smaller width or size than the gas inlet part 122.

A plurality of nozzle portions 123 may be formed along the gas inlet portion 122 extending in a circular shape. In addition, the plurality of nozzle units 123 are disposed to be spaced apart from each other.

The nozzle part 123 includes an inlet part 123a connected to the gas inlet part 122 and an outlet part 123b connected to an inner circumferential surface of the cylinder body 121. The nozzle part 123 is formed to have a predetermined length from the inlet part 123a toward the outlet part 123b.

The depth and width of the plurality of gas inlets 122 and the length of the nozzle part 123 are determined by the stiffness of the cylinder 120, the amount of the third filter 330, or the nozzle part 123 In consideration of the size of the pressure drop of the refrigerant passing through ), it can be determined as an appropriate size.

For example, when the depth and width of the plurality of gas inlets 122 are too large or the length of the nozzle unit 123 is too small, the stiffness of the cylinder 120 may be weakened.

On the other hand, if the depth and width of the plurality of gas inlets 122 are too small, the amount of the third filter 330 that can be installed in the gas inlet 122 may be too small.

In addition, when the length of the nozzle part 123 is too large, the pressure drop of the refrigerant passing through the nozzle part 123 becomes too large, so that a sufficient function as a gas bearing cannot be performed.

The diameter of the inlet part 123a of the nozzle part 123 is larger than the diameter of the outlet part 123b.

In detail, when the diameter of the nozzle part 123 is too large, the amount of the refrigerant flowing into the nozzle part 123 among the high-pressure gas refrigerant discharged through the discharge valve 161 becomes too large, resulting in a loss of flow rate of the compressor. There is a problem that becomes large.

On the other hand, when the diameter of the nozzle part 123 is too small, the pressure drop in the nozzle part 123 increases, and there is a problem that the performance as a gas bearing decreases.

Therefore, in this embodiment, the diameter of the inlet part 123a of the nozzle part 123 is relatively large to reduce the pressure drop of the refrigerant flowing into the nozzle part 123, and the diameter of the outlet part 123b By forming relatively small, it is characterized in that it is possible to adjust the inflow amount of the gas bearing through the nozzle unit 123 to a predetermined value or less.

A third filter 330 is installed in the plurality of gas inlets 122. The refrigerant flowing toward the inner circumferential surface of the cylinder 120 may be filtered by the third filter 330.

In detail, the third filter 330 blocks foreign substances of a predetermined size or more from flowing into the cylinder 120 and absorbs oil contained in the refrigerant. Here, the predetermined size may be 1 μm.

The third filter 330 includes a thread wound around the gas inlet 122. In detail, the thread is made of polyethylene terephthalate (PET) material and may have a predetermined thickness or diameter.

The thickness or diameter of the thread may be determined to be an appropriate value in consideration of the strength of the thread. If the thickness or diameter of the thread is too small, the strength of the thread becomes too weak and can be easily broken. If the thickness or diameter of the thread is too large, the thread is wound. In this case, there is a problem in that the air gap in the gas inlet 122 is too large to reduce the filtering effect of foreign matter.

For example, the thickness or diameter of the thread is formed in a unit of several hundred μm, and the thread may be formed by combining a number of spun threads in the unit of several tens of μm.

The thread is wound a plurality of times and the end thereof is configured to be fixed with a knot. The number of times the thread is wound may be appropriately selected in consideration of the degree of pressure drop of the gas refrigerant and the filtering effect of foreign matter. If the number of windings is too large, the pressure drop of the gas refrigerant becomes too large, and if the number of windings is too small, filtering of foreign matter may not be performed well.

In addition, the tension force wound around the thread is formed in an appropriate size in consideration of the degree of deformation of the cylinder 120 and the fixing force of the thread. If the tension is too large, deformation of the cylinder 120 may be caused, and if the tension is too small, the thread may not be well fixed to the gas inlet 122.

17 is a cross-sectional view showing a combination of a frame and a cylinder according to an embodiment, and FIG. 18 is an enlarged view of “B” of FIG. 17.

Referring to FIGS. 17 and 18, the linear compressor 100 according to the embodiment of the present invention includes a sealing pocket 370 in communication with the third flow path 430 and in which a sealing member 350 is installed. .

The sealing pocket 370 is a space in which the sealing member 350 can be installed, and is formed between the inner circumferential surface of the frame body 111 and the outer circumferential surface of the cylinder body 121. In addition, the sealing pocket 370 may be formed at the rear portion of the frame 110 and the cylinder 120. Based on the flow direction of the refrigerant, the flow cross-sectional area of the sealing pocket 370 is larger than the flow cross-sectional area of the third flow path 430.

In detail, the rear portion of the frame body 111 includes a pocket forming portion 112 configured to be recessed radially outward from the inner circumferential surface of the frame body 111. The pocket forming part 112 forms at least one surface of the sealing pocket 370.

In addition, the frame body 111 further includes a second inclined portion 113 extending obliquely in a rearward inward direction from the pocket forming portion 112.

The cylinder body 121 includes a first inclined portion 128 for forming the sealing pocket 370. The first inclined portion 128 constitutes at least one surface of the sealing pocket 370.

The first inclined portion 128 extends obliquely rearwardly from the first body end 121a of the cylinder body 121. In addition, the first inclined portion 128 may extend from an inner side of the pocket forming portion 112 to a point corresponding to an inner side of the second inclined portion 113.

Due to the recessed structure of the pocket forming part 112 and the inclined structure of the first inclined part 128, the height in the radial direction of the sealing pocket 370 may be larger than the diameter of the sealing member 350. have. In addition, a length of the sealing pocket 370 in the axial direction may be larger than a diameter of the sealing member 350.

That is, the sealing pocket 370 may have a size such that the sealing member 350 is movable without interfering with the frame body 111 or the cylinder body 121.

On the other hand, the distance or distance of the spaced space between the rear portion of the first inclined portion 128 and the rear portion of the second inclined portion 113 is formed smaller than the diameter of the sealing member 350. Therefore, during the operation of the linear compressor 100, when the refrigerant flows backward along the third flow path 430, the sealing member 350 moves backward by the pressure of the refrigerant, and the spaced space Will be sealed.

In this way, since the sealing member 350 is interposed between the cylinder 120 and the frame 110 to seal the third flow path 430, the refrigerant in the third flow path 430 ) Can be prevented from leaking to the outside.

In addition, the sealing member 350 is provided to be movable in the sealing pocket 370, and when a refrigerant flow occurs in the third flow path 430 when the compressor is driven, the sealing member 350 Since the cylinder 120 and the frame 110 are pressed, deformation of the cylinder 120 due to the pressing force of the sealing member 350 can be prevented.

Hereinafter, the flow of the refrigerant during the operation of the linear compressor will be described.

19 is a cross-sectional view showing the flow of refrigerant in the linear compressor according to the embodiment, FIG. 20 is a view showing the flow of the refrigerant discharged from the compression chamber according to the embodiment in the first and second flow paths, and FIG. 21 is It is a diagram showing the flow of refrigerant in the 3 flow path.

First, with reference to FIG. 19, the flow of the refrigerant in the linear compressor will be briefly described.

Referring to FIG. 19, the refrigerant flows into the shell 101 through the suction part 104 and flows into the suction muffler 150 through the suction guide part 155.

Then, the refrigerant flows into the second muffler 153 via the first muffler 151 of the suction muffler 150 and flows into the piston 130. In this process, the suction noise of the refrigerant can be reduced.

Meanwhile, the refrigerant may filter foreign matter having a predetermined size (25 μm) or more while passing through the first filter 310 provided to the suction muffler 150.

When the suction valve 135 is opened, the refrigerant passing through the suction muffler 150 and present in the piston 130 is sucked into the compression space P through the suction hole 133.

When the refrigerant pressure in the compression space P is equal to or higher than the discharge pressure, the discharge valve 161 is opened, and the refrigerant is discharged to the discharge space of the discharge cover 160 through the opened discharge valve 161. In detail, the discharge valve 161 moves forward and is spaced apart from the front surface of the cylinder 120, and in this process, the valve spring 162 is elastically deformed forward. In addition, the stopper 163 limits the amount of deformation of the valve spring 162 to a certain degree.

The refrigerant discharged to the discharge space of the discharge cover 160 flows to the discharge part 105 through the loop pipe 165 coupled to the discharge cover 160 and is discharged to the outside of the compressor 100.

Meanwhile, at least some of the refrigerants present in the discharge space of the discharge cover 160 are spaces between the cylinder 120 and the frame 110, that is, the first flow path 410 and the second flow path 420 ) Will flow. In addition, the refrigerant may be filtered by the second filter 320 while flowing through the first flow path 410 or the second flow path 420.

In addition, the filtered refrigerant flows toward the outer circumferential surface of the cylinder body 121 through the third flow path 430, and at least a portion of the refrigerant flows into the plurality of gas inlet portions 122 formed in the cylinder body 121. do. The refrigerant introduced into the gas inlet 122 is filtered by the third filter 330 and is introduced into the cylinder 120 through the nozzle unit 123.

The refrigerant introduced into the cylinder 120 is located between the inner circumferential surface of the cylinder 120 and the outer circumferential surface of the piston 130, and acts to separate the piston 130 from the inner circumferential surface of the cylinder 120 Do (gas bearing).

In this way, the high-pressure gas refrigerant is bypassed into the interior of the cylinder 120 to provide a floating pressure to the reciprocating piston 130 and act as a bearing at the interface between the cylinder and the piston. Accordingly, it is possible to reduce the wear between the piston 130 and the cylinder 120. In addition, by not using oil for bearings, even if the compressor 100 is operated at high speed, friction loss due to oil may not be generated.

The refrigerant flowing into the inner surface of the cylinder and the outer surface of the piston is a compressed refrigerant and has a condensing pressure (Pd). On the other hand, the space into which the refrigerant is introduced has an evaporation pressure (Ps) before being compressed. Accordingly, as the difference between the condensation pressure and the evaporation pressure increases, the piston 130 may float from the inner surface of the cylinder 120 with a greater force.

In addition, by providing a plurality of filters on the path of the refrigerant flowing inside the compressor 100, foreign matter contained in the refrigerant can be removed, and accordingly, the reliability of the refrigerant serving as a gas bearing can be improved. Accordingly, it is possible to prevent the occurrence of wear on the piston 130 or the cylinder 120 due to foreign substances contained in the refrigerant.

In addition, by removing the oil contained in the refrigerant by the plurality of filters, it is possible to prevent the occurrence of friction loss due to the oil. Since the first filter 310, the second filter 320, and the third filter 330 filter refrigerant to serve as a gas bearing, they may be collectively referred to as a “refrigerant filter device”.

Meanwhile, the refrigerant flowing through the third flow path 430 acts on the sealing member 350. In other words, the pressure of the refrigerant is applied to the sealing member 350, the sealing member 350 from the sealing pocket 370, the first inclined portion 128 of the cylinder 120 and the frame 110 It moves to the point between the second inclined parts 113 of.

In addition, the sealing member 350 is in close contact with the cylinder 120 and the frame 110, a spaced apart space between the cylinder 120 and the frame 110, for example, the first inclined portion 128 and The space between the second inclined portions 113 is sealed. Accordingly, it is possible to prevent the refrigerant in the third flow path 430 from leaking to the outside through the spaced apart space between the cylinder 120 and the frame 110.

On the other hand, when the driving of the linear compressor 100 is stopped, since the pressure of the refrigerant acting on the sealing member 350 is released, the adhesion between the sealing member 350 and the cylinder 120 and the frame 110 This weakens. As a result, the sealing member 350 is freely movable within the sealing pocket 220, for example, in a state spaced apart from the first inclined portion 128 and the second inclined portion 113 (dotted line mark). ).

According to this action, the sealing member 350 is in close contact with the cylinder 120 and the frame 110 only when the compressor 100 is driven to perform sealing of the third flow path 430, so that the sealing The force applied to the cylinder 120 from the member 350 may be reduced. Accordingly, deformation of the cylinder 120 can be prevented.

In addition, since the sealing member 350 may be placed in a movable state in the sealing pocket 370, interference of the sealing member 350 can be prevented when assembling the cylinder 120 and the frame 110. There will be. As a result, assembly of the cylinder 120 and the frame 110 may be facilitated.

As described above, by using the non-azeotropic mixed refrigerant, the floating pressure of the piston can be increased, and the action of the gas bearing can be smoothly performed. Since it is possible to provide a minimum floating pressure for the floating of the piston even in a low cooling operation section, the reliability of the linear compressor can be improved. For example, as the floating pressure increases and the friction between the contact surface between the cylinder and the piston decreases, it may be advantageous to increase the efficiency of the linear compressor through mechanical design such as a gas discharge hole and a filter.

In a refrigeration apparatus in which the continuous operation mode in which the low cooling power operation is frequently performed is performed, the effect can be maximized by applying the non-azeotropic mixed refrigerant to the oilless linear compressor.

On the other hand, the oilless linear compressor can reduce evaporation and combustion by heating oil due to friction between the cylinder and the piston when driving with low cooling power for a long time for the continuous operation mode, compared to a linear compressor using oil. There are also advantages.

In the above description, a linear compressor is preferably illustrated as a compressor. However, the present invention is not limited thereto, and may be preferably applied to all compressors in which oil circulation and air bearings are performed due to a pressure difference between the refrigerant when the gaseous refrigerant is compressed.

According to the present invention, since the non-azeotropic mixed refrigerant is used, it is possible to improve durability by preventing abrasion of parts due to the low evaporation temperature of the refrigerant in the continuous operation mode. In addition, it is possible to smoothly operate oil circulation and air bearings for lubrication essential for the operation of the compressor.

1: freshness of air
2: Freshness of non-azeotropic mixed refrigerant

Claims (20)

A compressor that is operated in a continuous operation mode to compress a non-azeotropic mixed refrigerant;
A condenser for condensing the refrigerant compressed by the compressor;
An expander for expanding the refrigerant condensed in the condenser; And
An evaporator for evaporating the refrigerant expanded in the expander is included,
A refrigeration apparatus using a non-azeotropic mixed refrigerant having a pressure difference (ΔP) of the non-azeotropic mixed refrigerant having a certain value included in the range of 340 kPa <ΔP <624.7 kPa. The method of claim 1,
The condensing pressure (Pd) of the non-azeotropic mixed refrigerant is a refrigeration apparatus using a non-azeotropic mixed refrigerant having a certain value within the range of 393.4 kPa <Pd <745.3 kPa. The method of claim 1,
A refrigeration apparatus using a non-azeotropic mixed refrigerant having a certain value included in the range of 53.5kPa <Ps <120.5kPa of the evaporation pressure (Ps) of the non-azeotropic mixed refrigerant. The method of claim 1,
In the continuous operation mode, the compressor is a refrigeration apparatus using a non-azeotropic mixed refrigerant that operates even in a temperature range where the internal temperature is satisfactory. The method of claim 1,
The compressor is a refrigeration device using a non-azeotropic mixed refrigerant which is a linear compressor. The method of claim 5,
In the linear compressor,
A shell provided with a suction unit;
A cylinder provided inside the shell and forming a compression space for a refrigerant;
A frame coupled to the outside of the cylinder;
A piston provided so as to reciprocate in the axial direction inside the cylinder;
A discharge valve that is movably coupled to the cylinder and selectively discharges the refrigerant compressed in the compression space of the refrigerant; And
A refrigeration apparatus using a non-azeotropic mixed refrigerant extending into a space between the cylinder and the frame and including a flow path through which at least some of the refrigerant discharged from the discharge valve flows. The method of claim 6,
In the cylinder,
A cylinder body in which the nozzle part is formed; And
A refrigeration apparatus using a non-azeotropic mixed refrigerant including a cylinder flange portion extending radially outward from the cylinder body. The method of claim 7,
In the frame,
A frame body surrounding the cylinder body;
A depression into which the cylinder flange portion is inserted; And
A refrigeration apparatus using a non-azeotropic mixed refrigerant including a seating portion facing the seating surface of the cylinder flange portion. The method of claim 8,
In the flow path,
A refrigeration apparatus using a non-azeotropic mixed refrigerant including a first flow path formed between an outer circumferential surface of the cylinder flange portion and an inner circumferential surface of the depression. The method of claim 8,
In the flow path,
A refrigeration apparatus using a non-azeotropic mixed refrigerant including a second flow path formed between the seating surface of the cylinder flange portion and the seating portion of the frame. The method of claim 8,
In the flow path,
A refrigeration apparatus using a non-azeotropic mixed refrigerant including a third flow path extending from the second flow path to a space between an outer circumferential surface of the cylinder body and an inner circumferential surface of the frame body. The method of claim 5,
In the linear compressor,
A cylinder in which a cylinder step portion is formed on an inner circumferential surface;
A piston formed on an outer circumferential surface of the cylinder so as to be linearly reciprocated, and forming a low pressure between the stepped portions of the cylinder when moving in one direction and forming a high pressure between the stepped portions of the cylinder when moving in the other direction;
An oil suction passage formed to allow oil to flow between the cylinder stepped portion and the piston stepped portion; And
A refrigeration apparatus using a non-azeotropic mixed refrigerant including an oil discharge passage formed to allow oil between the cylinder stepped portion and the piston stepped portion to be discharged to the outside of the cylinder. The method of claim 1,
The non-azeotropic refrigerant is composed of the first hydrocarbon and the second hydrocarbon, the first hydrocarbon is isobutane, and the second hydrocarbon is propane. The method of claim 13,
The isobutane is a refrigeration apparatus using a non-azeotropic mixed refrigerant provided in a weight ratio of 76% ≤ isobutane ≤ 87%. A linear compressor for compressing a non-azeotropic mixed refrigerant;
A condenser for condensing the non-azeotropic mixed refrigerant compressed by the compressor;
An expander for expanding the non-azeotropic mixed refrigerant condensed in the condenser; And
An evaporator for evaporating the non-azeotropic mixed refrigerant expanded in the expander is included,
A refrigeration apparatus using a non-azeotropic mixed refrigerant having a pressure difference (ΔP) of the non-azeotropic mixed refrigerant of 340 kPa <ΔP <624.7 kPa. The method of claim 15,
In the linear compressor,
-Piston for reciprocating motion; And
-A cylinder guiding the piston is included,
A refrigeration apparatus using a non-azeotropic mixed refrigerant compressed by the piston to guide the high-pressure non-azeotropic refrigerant to the inner surface of the cylinder and to float the outer surface of the piston from the inner surface of the cylinder. The method of claim 15,
The non-azeotropic mixed refrigerant contains at least two hydrocarbons, and in the at least two hydrocarbons,
-At least one first hydrocarbon selected from the phase group having an evaporation temperature of -12 degrees or more at 1 bar; And
-At least one second hydrocarbon selected from the middle group having an evaporation temperature of -50 degrees or more and less than -12 degrees at 1 bar is included,
A refrigeration device that uses a non-azeotropic mixed refrigerant with a temperature gradient of more than 4 degrees Celsius. The method of claim 17,
A refrigeration apparatus using a non-azeotropic mixed refrigerant having a weight ratio of the first hydrocarbon of 50% or more. The method of claim 15,
In the linear compressor,
-A piston for compressing the non-azeotropic mixed refrigerant by reciprocating motion; And
-A cylinder guiding the piston is included,
A refrigeration apparatus using a non-azeotropic mixed refrigerant having oil pumped by a pressure difference between the non-azeotropic mixed refrigerant on a contact surface between the piston and the cylinder. A compressor controlled to compress the non-azeotropic mixed refrigerant;
A condenser for condensing the refrigerant compressed by the compressor;
An expander for expanding the refrigerant condensed in the condenser; And
An evaporator for providing cool air in the chamber by evaporating the refrigerant expanded in the expander is included,
The compressor,
-It is operated by selecting intermittent operation mode and continuous operation mode,
-In the continuous operation mode, a refrigeration apparatus using a non-azeotropic mixed refrigerant in which the compressor is controlled so that the compressor is continuously operated even when the temperature in the container is within the range of the target temperature.

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